1. Field of the Invention
The present invention generally relates to automotive air conditioning systems. More specifically, this invention is directed to a receiver-dryer for use in an automotive air conditioning system wherein the receiver-dryer includes unique features for improving the efficiency of the separation of a gas phase from a liquid phase of a refrigerant fluid and for redirection of the liquid phase so as to improve sub-cooling of the refrigerant through the receiver-dryer and a condenser.
2. Description of the Related Art
Air-conditioning systems for motor vehicles are well known. FIG. 5 illustrates an example of a typical air-conditioning system 10, which essentially includes a compressor 12, a condenser 14, a thermal expansion valve 16, an evaporator 18, a refrigerant line 20 connecting the aforementioned components together, and a refrigerant fluid flowing therethrough (as represented by the various arrows). It is also known to provide a receiver-dryer 22 in a refrigeration circuit between the condenser 14 and the thermal expansion valve 16 to remove particulates and moisture from the refrigerant fluid and thereby protect the downstream components.
At the beginning of a refrigeration cycle, an upstream side 24 of the compressor 12 receives a gaseous phase of the refrigerant fluid. Powered by an engine of the motor vehicle (not shown) via a belt drive 26 and clutch 28 or electrically driven system, the compressor 12 compresses the refrigerant fluid to increase the temperature and pressure to create a superheated vapor and to pump the refrigerant downstream through the refrigerant line 20 to the condenser 14.
Within the condenser 14, the superheated refrigerant fluid changes from its gaseous phase to a mostly liquid phase. The superheated vapor of the refrigerant fluid flows through interior passages 30 of the condenser 14 while ambient air flows over exterior surfaces 32 and cooling fins 34 of the condenser 14. The superheated vapor is much hotter than the ambient air. Thus, the heat of the superheated vapor is given off to the surrounding ambient air flowing over the exterior surfaces 32 and cooling fins 34 of the condenser 14, thereby cooling the refrigerant fluid in accord with heat transfer principles. As the refrigerant fluid continues to flow through the condenser 14 and lose more heat to the surrounding ambient air, it begins to condense from its gaseous phase into a liquid phase. Eventually, the refrigerant fluid exits the condenser 14, mostly in a liquid phase (X) but typically including some gaseous portion, and flows downstream through the refrigerant line 21, and enters the receiver-dryer 22.
The receiver-dryer 22 includes an adsorbent unit 36 therein for dehydrating or removing water from the refrigerant fluid. The receiver-dryer 22 includes an outlet line 38 having a pickup end 40 disposed in a lower region 42 for communicating only liquid phase, and not gaseous phase, refrigerant out of the receiver-dryer 22 and downstream to the thermal expansion valve 16.
The thermal expansion valve 16 “expands” the refrigerant fluid so as to suddenly reduce the pressure of the refrigerant fluid. This sudden reduction in pressure causes the refrigerant fluid to be sprayed through the refrigerant line 20 downstream to the evaporator 18.
Within the evaporator 18, the evaporation process extracts the required evaporator heat from an incoming stream of fresh or recirculating interior air, thereby cooling the air. The now latent heat of liquid fluid phase of the refrigerant fluid changes back into a gaseous phase as a result of the heat received from the fresh or recirculating interior air. While the now relatively cool refrigerant fluid flows through interior passages (not shown) of the evaporator 18, relatively hot ambient air flows over exterior surfaces (not shown) of the evaporator 18, in similar fashion as the condenser 14. The evaporator 18 cools the hot moist ambient air because the humidity or water vapor in the hot ambient air collects or condenses on the exterior surfaces of the evaporator 18. The evaporator 18 also dehumidifies the hot moist ambient air because the moist ambient air is given off to the relatively cold refrigerant flowing through the evaporator 18, thereby warming the refrigerant fluid and cooling the air flowing over the exterior surfaces of the evaporator 18. Thus, a supply of cool, dry, dehumidified air flows away from the evaporator 18 and into a passenger compartment of the motor vehicle (not shown), while the heated gaseous refrigerant flows out of the interior passages of the evaporator 18, through the refrigerant line 20 downstream back to the compressor 12 where the refrigeration cycle repeats.
Referring to prior art FIGS. 5 and 6, there is shown a pressure vs. enthalpy diagram of the prior art refrigeration cycle with pressure depicted along the ordinate and enthalpy depicted along the abscissa. Schematic points O, A, D, and F of FIG. 5 are graphically represented in FIG. 6 as points O, A, D, and F of the refrigeration cycle. In general, path O–A represents the compression stage of the refrigeration cycle, path A–D represents the condensing stage, path D–F represents the expansion stage, and path F–O represents the evaporation stage of the refrigeration cycle. Point B represents the transition point at which the refrigerant condenses from a superheated vapor to a saturated vapor. Point C represents the transition point at which the refrigerant further condenses from a liquid-vapor mixture to a saturated liquid.
In prior art air-conditioning systems, under vehicle usage conditions there may—or may not—be sub-cooling at the output side (range X—in FIG. 6, B–C) of the condenser (14 in FIG. 5), depending upon the state of the refrigerant fluid due to various vehicle performance variables. In other words, and referring to FIG. 6, range X represents the variable nature of the refrigerant fluid temperature at the downstream or output side of the condenser 14 at range X in FIG. 5 and Y1 represents the sub-cooling of prior art refrigeration cycle. Whereas point A is well defined and fixed at the location on the pressure vs. enthalpy diagram as shown, range X is not so well defined and varies along the condenser path A–D of the pressure vs. enthalpy diagram depending upon the vehicle performance variables of vehicle speed and load on the air-conditioning system. The slower the vehicle speed, or at idle condition and, the higher the load on the air-conditioning system, the sub-cooling range Y1 diminishes and may approach zero. Under these conditions, the refrigeration cycle looses sub-cooling capability and operates only in the “X” range. Likewise, point D is dependent upon the amount of sub-cooling that can be performed on the refrigerant beyond point C. In other words, point D is incrementally dependent upon the cooling load and quantity of ambient air flow when the air conditioning system is properly charged with refrigerant.
Referring to FIG. 6, the amount of heat (Q) that can be removed by the condenser (14) is represented by the equation Q=MR134a* (h2−h1). MR134a is the variable mass flow for R134a refrigerant while h2 is the enthalpy at the beginning of the refrigerant entering into the condenser, 14 and h1 is the enthalpy at the receiver dryer outlet D. Assuming a constant mass flow, the greater the range in enthalpy that the air-conditioning system can produce, the greater the heat that can be removed.
More recent advancements in automotive refrigeration suggest structurally integrating a receiver-dryer with a condenser. For example, U.S. Pat. No. 5,927,102 to Matsuo et al. teaches a receiver that is integrally mounted to a condenser in such a manner as to maintain a constant sub-cool temperature. The '102 patent discloses the condenser as including a pair of opposed and vertically extending first and second header tanks and a core composed of a plurality of tubes extending between the header tanks in a generally horizontal fashion. At the top of the first header tank, an inlet joint is disposed into which superheated refrigerant from the compressor flows. At the bottom of the second header tank, an outlet joint is disposed out of which substantially condensed refrigerant flows. Inner spaces of the header tanks are divided by separators into an upper space into which the superheated refrigerant flows and a lower space into which flows refrigerant cooled down in the core. The receiver is mounted to the condenser in fluidic communication between the upper and lower spaces of the condenser. More specifically, the receiver-dryer is mounted to the condenser such that the receiver does not overlap with the upper space in order to minimize heat transfer from the incoming superheated refrigerant to the refrigerant fluid collected in the receiver, thereby minimizing evaporation of the refrigerant fluid. Accordingly, a “whole” space of the receiver can be reserved for adding make up refrigerant to compensate for loss of refrigerant due to leakage, while maintaining a constant sub-cool temperature.
From the above, it can be appreciated that receiver-dryers of the prior art are not fully optimized. For example, while the '102 patent does teach passive stabilization of the sub-cooling temperature of the condenser, it does not teach active optimization of sub-cooling of the condenser. In other words, the '102 patent focuses on passively avoiding evaporation of the liquid phase of the refrigerant fluid within the condenser, rather than actively maximizing condensing of the gas phase into the liquid phase. Moreover, the performance of the prior art receiver-dryer of FIGS. 5 and 6 is excessively dependent upon vehicle operating conditions and air conditioning demand. Thus, there remains a need for an integrated receiver-dryer that is less dependent upon vehicle operating conditions and air conditioning demand, and that not only minimizes evaporation of a liquid phase therein, but also maximizes the liquid phase so as to return relatively more liquid phase to the condenser for additional sub-cooling, thereby enabling the condenser to consistently output 100% sub-cooled liquid phase refrigerant.